Turbochargers deliver air, at greater density than would be possible in the normally aspirated configuration, to the engine intake, allowing more fuel to be combusted, thus boosting the engine's horsepower without significantly increasing engine weight. This can enable the use of a smaller turbocharged engine, replacing a normally aspirated engine of a larger physical size, thus reducing the mass and aerodynamic frontal area of the vehicle.
Turbochargers are a type of forced induction system which use the exhaust flow entering the turbine housing from the engine exhaust manifold to drive a turbine wheel (51) which is located in the turbine housing. The turbine wheel is solidly affixed to a shaft to become the shaft and wheel assembly. A compressor wheel (20), is mounted to the stub shaft (56) end of the shaft and wheel and held in position by the clamp load from a compressor nut (29). The primary function of the turbine wheel is extracting rotational power from the exhaust gas to drive the compressor.
The compressor stage consists of a wheel (20) and its housing. Filtered air is drawn axially into the inlet of the compressor cover by the rotation of the compressor wheel. The power input by the turbine stage to the shaft and wheel drives the compressor wheel to produce a combination of static pressure with some residual kinetic energy and heat. The pressurized gas exits the compressor cover through the compressor discharge and is delivered, usually via an intercooler, to the engine intake.
In one aspect of compressor stage performance, the efficiency of the compressor stage is influenced by the clearances between the compressor wheel contour (28) and the matching contour in the compressor cover. The closer the compressor wheel contour is to the compressor cover contour, the higher the efficiency of the stage. In a typical compressor stage with a 76 mm compressor wheel, the tip clearance is in the regime of from 0.31 mm to 0.38 mm. The closer the wheel is to the cover, the higher the chance of a compressor wheel rub; so, there must exist a compromise between improving efficiency and improving durability.
To the naked eye, the nose of the compressor wheel in a typical turbocharger appears to rotate about the geometric longitudinal axis of the bearing housing; however, when viewed as a track on an x,y-oscilloscope, the nose of the compressor wheel describes orbits of various shapes. The average centroid of the orbits is close to, but not exactly centered on, the geometric longitudinal axis of the turbocharger. The geometric axis (100) is shown in FIG. 1, of the turbocharger.
The dynamic excursions taken by the shaft and wheels are attributed to a number of factors including: the unbalance of the rotating assembly; the excitation of the pedestal (i.e., the engine and exhaust manifold); and the low speed excitation from the vehicle's interface with the ground.
The net effect of these excursions taken by the shaft and wheels is that the design of the typical turbocharger has clearances far greater than those desired for aerodynamic efficiency levels.
The typical turbocharger is fed with oil from the engine. This oil, at a pressure typically equal to that of the engine, performs several functions. The oil is delivered to both sides of the journal bearings to provide a double hydrodynamic squeeze film, the pressures of which exert reactionary forces of the shaft on the inner diameter (I.D.) of the bearing and of the outer diameter (O.D.) of the bearing on the bearing housing bore. The oil films provide attenuation of the reactionary forces to reduce the amplitude of the excursions of the shaft. The oil also functions to remove heat from the turbocharger.
A typical turbocharger design has two adjacent bearing systems: one at the compressor-end of the bearing housing; and one at the turbine-end of the bearing housing. Each system has two interfaces: the interface of the rotating shaft on the I.D. of the floating bearing, and the interface of the O.D. of the floating bearing on the fixed bore of the bearing housing.
The stiffness and damping capacities of the typical turbocharger double hydrodynamic squeeze film bearings are a compromise between: the thickness of the oil film generated by the rotational speed of the bearing elements, the clearance between said elements, and the oil flow limitations due to the propensity of turbochargers to pass oil through the piston ring seals at either end of the shaft.
The problems of loss of efficiency due to excessive clearances between wheels and housings include: high oil flow rates for bearing support, bearing damping, heat transfer, and power losses, and these problems are solved by the use of rolling element bearings (REB) to support and locate the rotating assembly in the turbocharger.
FIG. 1 depicts a typical turbocharger double hydrodynamic squeeze film bearing configuration. In this configuration, pressurized oil is received to the bearing housing (3) though an oil inlet (80) from the engine. The oil is pressure-fed through the oil galleries (82 and 83) to the bearing housing journal bearing bore. For both the turbine-end and compressor-end bearings (30), the oil flow is delivered to the shaft and wheel journal bearing zones where the oil is distributed around the shaft to generate an oil film between the shaft surface (52) and the inner bore of the floating journal bearings (30). On the outside of the journal bearings (30), a like oil film is generated by the rotation of the journal bearing against the bearing housing journal bearing bore. Once through the journal and thrust bearings the oil exits the bearing housing via the oil drain (85) at the base of the bearing housing, and is returned to the crankcase of the engine.
In the typical turbocharger depicted in FIG. 1, the thrust bearing (19) is also a hydrodynamic or fluid film type of bearing. In this configuration, the stationary thrust bearing is fed oil from the oil gallery (81) to feed a ramp and pad design of the bearing. The oil is driven into a wedge shape by the relative motion of the thrust washer (40) and the opposing face of the flinger (44), which is mounted to the shaft, against the static thrust ramp and pad. This bearing controls the axial position of the rotating assembly.
One method for increasing the efficiency of the turbocharger is the adoption of rolling element bearings (REBs) to support the rotating assembly. Rolling element bearings can be divided into two general types. The first type uses a pair of typical REB assemblies. Each REB assembly, in this case, consists of an outer race, the balls or roller elements, an inner race, a cage, and seals. This pair of REB assemblies can be pressed or shrunk into a sleeve, i.e., an outer cylindrical housing with oil galleries and locations for the REB assemblies, to produce the REB cartridge. In the second type, the sleeve is omitted, and the outer race of the REB assembly defines the outer diameter of the REB cartridge. Unless otherwise indicated, the term “REB” used herein will refer to the REB cartridge.
As seen in FIG. 2 REBs typically have an inner race (65), which is mounted to the shaft and wheel journal surface (52). Assembled to the inner race, or races (65, 65C and 65T) are a set of rolling elements connecting the inner race to the outer race (64) (FIG. 6). The outer race is mounted within the bore (71) in the bearing housing (3). Since rolling element bearings do not require as much oil as do typical turbocharger journal bearings, an oil restrictor (86) is fitted to the oil inlet (80) to restrict the flow REBs.
There are several improvements that come with the adoption of rolling element bearing turbochargers. There is an improvement in transient response due to the reduction in power losses, especially at low turbocharger RPM, of the REB system over the typical turbocharger bearing system. The power losses in REB systems are from 5 to 10 times less than those for typical sleeve-type turbocharger bearing systems. REB systems can support much greater thrust loads than can typical turbocharger bearing systems making the thrust component more robust. Since typical ramp and pad thrust bearings require a large percentage of the oil flow delivered to the turbocharger, and REB systems require less oil flow (than a typical turbocharger bearing system), then less oil flow is required for a REB system with the positive consequence that there is less propensity for oil passage to the compressor or turbine stages where that oil can poison the catalyst.
Although ball bearing systems provide these efficiency and transient performance gains, the damping capacity of ball bearings is not as good as that of the typical turbocharger double hydrodynamic squeeze film bearings. For ease of assembly, the ball bearings are retained in a steel REB cartridge, which is suspended within the bearing housing by an oil film between the O.D. of the cartridge (172) and the I.D. of the bearing housing bore (71). The oil is used for damping of shaft critical events and for lubrication of the bearings. With this design it is critical that the bearing cartridge is not in a metal-to-metal contact with the bearing housing bore as the damping function will be lost.
U.S. Pat. No. 5,145,334 (Gutknecht) and U.S. Pat. No. 7,214,037 (Mavrosakis) teach methods for the retention of a floating bearing cartridge in a bearing housing. A post secured in the bearing housing (e.g., restrictor (86)) restrains the bearing cartridge such that the post reacts against the axial and rotational forces, while allowing for otherwise unconstrained motion (float) of the bearing cartridge in the bearing housing. In U.S. Pat. No. 7,214,037, as shown in FIG. 4, a pin (460) received by an opening (444) of the housing (440) optionally aids in locating the cartridge azimuthally, with respect to the housing (440). A pin (72) similar to that of U.S. Pat. No. 5,145,334 is shown in FIG. 3 of the present application, locating a bore (68) in the cartridge and a bore (70) in the outer race to provide both axial and rotational constraint relative to the bearing housing. Both of these methods require machining through orifices in the bearing housing, and, furthermore, they require intricate assembly in an area not well-visible to the assembler, making both correct assembly and verification of the assembly of said pins difficult.
U.S. Pat. No. 7,214,037 teaches the use of a counter-bore (442, FIG. 4) and a plate (450) to control the axial loads exerted on the outer race of the bearing cartridge. The machining of this counter-bore requires accurate placement of the cutting tool, which must change direction from cutting a diametral surface to an abutment surface deep inside the bearing housing, without leaving too great an inclusive corner radius which might not allow the bearing cartridge to seat on the abutment. The execution of this process adds cost and complexity to the machining of the bearing housing.
So it can be seen that the current state of axial and rotational constraint of the REB cartridge is both costly and complex. A more cost and technically effective solution is needed.